Auto rack rail road cars are used to transport automobiles. Most often, although not always, they are used to transport finished automobiles from a factory or a port to a distribution center. Typically, auto-rack rail road cars are loaded in the “circus loading” manner, by driving vehicles into the cars from one end, and securing them in places with chocks, chains or straps. When the trip is completed, the chocks are removed, and the cars are driven out.
Automobile manufacturers would like to be able to have new cars driven into the auto-rack cars, and then to be held in place using the parking brake of the car alone, without the need for chocks or chains. At present the operating characteristics of auto-rack cars are not generally considered to be gentle enough to permit this do be done reliably. That is, a long standing concern has been the frequency of damage claims arising from high accelerations imposed on the lading during train operation. It has been suggested that the maximum design load condition of some automobile components occurs during the single journey of the automobile on the rail car.
Damage due to dynamic loading in the rail car may tend to arise principally in two ways. First, there are the longitudinal input loads transmitted through the draft gear due to train line action or shunting. Second, there are vertical, rocking and transverse dynamic responses of the rail road car to track perturbations as transmitted through the rail car suspension.
In this context, slack includes (a) the free slack in the couplers; and (b) the travel of the draft gear of successive rail road cars under the varying buff and draft loads. Slack run-out occurs, for example, as a train climbs a long upgrade, and all of the slack is taken out of the couplings as the train stretches. Once the train clears the crest, and begins its descent, the rail road cars at the end of the train may tend to accelerate downhill into the cars in front, closing up the slack. This slack run-in and run-out can result in significant longitudinal accelerations. These accelerations are transmitted to the automobiles carried in the auto-rack cars.
Historically, the need for slack was related, at least in part, to the difficulty of using a steam locomotive to “lift” (that is, move from a standing start) a long string of cars with journal bearings, particularly in cold weather. Steam engines were reciprocating piston engines whose output torque at the drive wheels varied as a function of crank angle. By contrast, presently operating diesel-electric locomotives are capable of producing high tractive effort from a standing start, without concern about crank angle or wheel angle. For practical purposes, presently available diesel-electric locomotives are capable of lifting a unit train of one type of cars having little or no slack.
Switching is another process having a long history. Two common types of switching are “flat switching” and “humping”. Humping involves running freight cars successively over a raised portion of track, and then allowing the car to run down-hill under gravity along various leads and sidings to couple with other cars as a train consist is assembled. For this type of operation the coupling speeds can be excessive, resulting in similarly excessive car body accelerations. For many types of rail road car, humping is now forbidden due to the probability of damaging the lading. An alternate form of switching is “flat switching” in which a locomotive is used to give a push to a rail road car, and then to send it rolling under its own inertia down a chosen siding to couple with another car. Particularly when done at night, the desirability of making sure that a good coupling is made tends to encourage rail yard personnel to make sure that the rail road cars are given an extra generous push. This often less than gentle habit tends to lead to rather high impact loads during coupling at impacts in the 5 m.p.h. (or higher) range. Forces can be particularly severe when there is an impact between a low density lading rail road car, such as an auto rack car, and a high density lading car (or string of cars) such as coal or grain cars.
Given this history, rail road car draft gear are designed to cope with slack run-out and slack run-in during train operation, and also to cope with the impact as cars are coupled together. Historically, common types of draft gear, such as that complying with, for example, AAR specification M-901-G, have been rated to withstand an impact at 5 m.p.h. (8 km/h) at a coupler force of 500,000 Lb. (roughly 2.2×106 N). Typically, these draft gear have a travel of 2¾ to 3¼ inches in buff before reaching the 500,000 Lb. load, and before “going solid”. The term “going solid” refers to the point at which the draft gear exhibits a steep increase in resistance to further displacement. If the impact is large enough to make the draft gear “go solid” then the force transmitted, and the corresponding acceleration imposed on the lading, increases sharply. While this may be acceptable for coal or grain, it is undesirably severe for more sensitive lading, such as automobiles or auto parts, paper, and other high value consumer goods such as household appliances.
Consequently, from the relatively early days of the automobile industry, there has been a history of development of longer travel draft gear to provide lading protection for relatively high value, low density lading, in particular automobiles and auto parts, but also farm machinery, or tractors, or highway trailers. Draft gear development has tended to be directed toward providing longer travel on impact to reduce the peak acceleration. In the development of sliding sills, and latterly, hydraulic end of car cushioning (EOCC) units, the same impact is accommodated over 10, 15, or 18 inches of travel. As a result, for example, by the end of the 1960's nearly all auto rack cars, and other types of special freight cars had EOCC units. Further, of the approximately 45,000 auto-rack cars in service in 1997, virtually all were equipped with end of car cushioning units. A discussion of the developments of couplers, draft gear and EOCC equipment is given the 1997 Car and Locomotive Cyclopedia (Simmons-Boardman Books, Inc., Omaha, 1997 ISBN 0-911382-20-8) at pp. 640-702. In summary, there has been a long development of long travel draft gear equipment to protect relatively fragile lading from end impact loads.
In light of the foregoing, it is counter-intuitive to employ short-travel, or ultra short travel, draft gear for carrying wheeled vehicles. However, by eliminating, or reducing, the accumulation of slack, the use of short travel buff gear may tend to reduce the relative longitudinal motion between adjacent rail road cars, and may tend to reduce the associated velocity differentials and accelerations between cars. The use of short travel, or ultra-short travel, buff gear also has the advantage of eliminating the need for relatively expensive, and relatively complicated EOCC units, and the fittings required to accommodate them. This may tend to permit savings both at the time of manufacture, and savings in maintenance during service.
Further, as noted above, given the availability of locomotives that develop continuous high torque from a standing start, it is possible to re-examine the issue of slack action from basic principles. The use of vehicle carrying rail road cars in unit trains that will not be subject to operation with other types of freight cars, that will not be subject to flat switching, and that may not be subject to switching at all when loaded, provides an opportunity to adopt a short travel, reduced slack coupling system throughout the train. The conventional approach has been to adopt end of car equipment with sufficient travel to cope with existing slack accumulation between cars. In doing so, the long travel end of car equipment has tended to add to the range of slack action in the train that is to be accommodated by the draft gear along the train. The opposite approach is to avoid a large accumulation of slack in the first place. If a large amount of slack is not allowed to build up along the train, then the need for long-travel draft gear and other end of car equipment is also reduced, or, preferably, eliminated.
One way to reduce slack action is to use fewer couplings. To that end, since articulated connectors are slackless, use of articulated rail road cars significantly reduces the slack action in the train. Some releasable couplings are still necessary, to permit the composition of a train to change, if desired. Further, it is necessary to be able to change out a car for repair or maintenance when required.
To reduce overall slack, it would be advantageous to adopt a reduced slack, or slackless, coupler, (as compared to AAR Type E). Although reduced slack AAR Type F couplers have been known since the 1950's, and slackless “tightlock” AAR Type H couplers became an adopted standard type on passenger equipment in 1947, AAR Type E couplers are still predominant. AAR Type H couplers are expensive, (and are used for passenger cars), as were the alternate standard Type CS controlled slack couplers. According to the 1997 Cyclopedia, supra, at p. 647 “Although it was anticipated at one time that the F type coupler might replace the E as the standard freight car coupler, the additional cost of the coupler and its components, and of the car structure required to accommodate it, have led to its being used primarily for special applications”. One “special application” for F type couplers is in tank cars, another is in rotary dump coal cars.
The difference between the nominal ⅜″ slack of a Type F coupler and the nominal 25/32″ slack of a Type E coupler may seem small in the context of EOCC equipped cars having 10, 15 or 18 inches of travel. By contrast, that difference, 13/32″, seems proportionately larger when viewed in the context of the approximately 11/16″ buff compression (at 700,000 lbs.) of Mini-BuffGear. It should be noted that there are many different styles of Type E and Type F couplers, whether short or long shank, whether having upper or lower shelves, as described in the Cyclopedia, supra. There is a Type E/F having a Type E coupler head and a Type F shank. There is a Type E50ARE knuckle which reduces slack from 25/32 to 20/32″. Type F herein is intended to include all variants of the Type F series, and Type E herein is intended to include all variants of the Type E series having 20/32″ of slack or more.
Another way to reduce slack action in the draft gear is to employ stiffer draft gear. Short travel draft gear are presently available. As noted above, most M-901-G draft gear have an official rating travel of 2¾″ to 3¼″ under a buff load of 500,000 Lbs. Mini-BuffGear, as produced by Miner Enterprises Inc., of 1200 State Street, Geneva Ill., appears to have a displacement of less than 0.7 inches at a buff load of over 700,000 lbs., and a dynamic load capacity of 1.25 million pounds at 1 inch travel. This is nearly an order of magnitude more stiff than some M-901-G draft gear. Miner indicates that this “special BuffGear gives drawbar equipped rail cars and trains improved lading protection and train handling”, and further, “[The resilience of the Mini-BuffGear] reduces the tendency of the draw bar to bind while negotiating curves. At the same time, the Mini-BuffGear retains a high pre-load to reduce slack action. Elimination of slack between coupler heads, plus Mini-Buff Gear's high pre-load and limited travel, provide ultralow slack coupling for multiple-unit well cars and drawbar connected groups of unit train coal cars.” Notably, unlike vehicle carrying rail cars, coal is unlikely to be damaged by the use of short travel draft gear.
In addition to M-901-G draft gear, and Mini-BuffGear, it is also possible to obtain draft gear having less than 1¾ inches of deflection at 400,000 Lbs., one type having about 1.6 inches of deflection at 400,000 Lbs. This is a significant difference from most M-901-G draft gear.
In terms of dynamic response through the trucks, there are a number of loading conditions to consider. First, there is a direct vertical response in the “vertical bounce” condition. This may typically arise when there is a track perturbation in both rails at the same point, such as at a level crossing or at a bridge or tunnel entrance where there may be a sharp discontinuity in track stiffness. A second “rocking” loading condition occurs when there are alternating track perturbations, typically such as used formerly to occur with staggered spacing of 39 ft rails. This phenomenon is less frequent given the widespread use of continuously welded rails, and the generally lower speeds, and hence lower dynamic forces, used for non-welded track. A third loading condition arises from elevational changes between the tracks, such as when entering curves in which case a truck may have a tendency to warp. A fourth loading condition arises from truck “hunting”, typically at higher speeds, where the conicity of the wheels tends not only to give the trucks a measure of self-steering ability, but tends also to cause the truck to oscillate transversely between the rails. During hunting, the trucks tend most often to deform in a parallelogram manner. Lateral perturbations in the rails sometimes arise where the rails widen or narrow slightly, or one rail is more worn than another, and so on.
There are both geometric and historic factors to consider related to these loading conditions. One is the near universal usage of the three-piece style of freight car truck in North America. While other types of truck are known, such as an H-frame truck or single axle fixed truck as used in Europe, the three piece truck has advantages that have made it overwhelmingly dominant in freight service in North America. First, it can carry greater loads than a fixed, single axle truck, and permits greater longitudinal truck spacing than a single axle truck. The three piece truck is simple. It employs only three main component elements, namely a truck bolster and a pair of side frames. The side frame castings are inexpensive relative to alternative H-frame designs. Manufacture of the side frame requires a relatively small mold as compared to an H-frame truck, and may tend to be less prone to molding defects. The three piece truck relies on a primary suspension in the form of a set of springs trapped in a “basket” between the truck bolster and the side frames. The three piece truck can operate in a wide range of environmental conditions, over a long period of time, with relatively little maintenance. When maintenance is required, the springs and axles can be changed out relatively easily. In terms of wheel load equalisation, a three piece truck uses one set of springs and the side frames pivot about the truck bolster ends in a manner like a walking beam. By contrast, an H frame truck requires both a primary suspension and secondary suspension at each of the wheels. In summary, the 1980 Car & Locomotive Cyclopedia, states at page 669 that the three piece truck offers “interchangeability, structural reliability and low first cost but does so at the price of mediocre ride quality and high cost in terms of car and track maintenance”. It would be desirable to retain many or all of these advantages while providing improved ride quality.
In terms of loading regimes, the first consideration is the natural frequency of the vertical bounce response. The static deflection from light car (empty) to maximum laded gross weight (full) of a rail car at the coupler must tend not to fall outside a given range, typically about 2 inches, if the couplers are to perform satisfactorily in interchange service. In addition, rail road car suspensions have a dynamic range in operation, including a reserve allowance.
In typical historical use, springs were chosen to suit the deflection under load of a full coal car, or a full grain car, or full loaded general purpose flat car. In each case, the design lading tended to be very heavy relative to the rail car weight. The live load for a 286,000 lbs., car may be of the order of five times the weight of the dead sprung load (i.e., the weight of the car including truck bolsters but less side frames, axles and wheels). Further, in these instances, the lading may not be particularly sensitive to abusive handling. That is, neither coal nor grain tends to be damaged badly by excessive vibration. In addition, coal and grain tend to have a relatively low value per unit weight. As a result these cars tend to have very stiff suspensions, with a dominant natural frequency in vertical bounce mode of about 2 Hz. when loaded, and about 4 to 6 Hz. when empty. Historically, much effort has been devoted to making freight cars light for two reasons. First, the weight to be back hauled empty is kept low, reducing the fuel cost of the backhaul. Second, when the ratio of lading to car weight increases, a higher proportion of hauling effort goes into hauling lading, as opposed to hauling the deadweight of the railcars themselves.
By contrast, an autorack car has the opposite loading profile. A two unit articulated autorack car as presently in service may have a light car weight of 165,000 lbs., and a lading weight when fully loaded of only 35-40,000 lbs. The lading typically has a high, or very high, ratio of value to weight. Generally, while coal may account for as much as 40% of all car loadings, it may generate only about 25% of freight revenues. By comparison, automobiles may account for only about 2% of car loadings, yet may account for about 10% of freight revenues. Similarly, unlike coal or grain, automobiles are relatively fragile, and hence more sensitive to a gentle (or a not so gentle) ride. As a relatively fragile, high value, high revenue form of lading, it may be desirable to incur a greater expense to obtain superior ride quality to that suitable for coal or grain.
Historically auto rack cars were made by building a rack structure on top of a general purpose flat car. As such, the resultant car was sprung for the flat car design loads. This might yield a vertical bounce natural frequency in the range of 3 Hz. It would be preferable for the rail car vertical bounce natural frequency to be on the order of 1.4 Hz or less. Since this natural frequency varies as the square root of the quotient obtained by dividing the spring rate of the suspension by the overall sprung mass, it is desirable to reduce the spring constant, to increase the mass, or both.
Deliberately increasing the mass of any kind of freight car is, itself, counter intuitive, since many years of effort has gone into reducing the weight of rail cars relative to the weight of the lading for the reasons noted above. One manufacturer, for example, advertises a light weight aluminium auto-rack car. However, given the high value and low density of the lading, adding weight may be reasonable to obtain a desired level of ride quality. Further, auto rack rail cars tend to be tall, long, and thin, with the upper deck loads carried at a relatively high location as measured from top of rail. A significant addition of weight at a low height relative to top of rail may also be beneficial in reducing the height of the center of gravity of the loaded car.
Decreasing the spring rate involves further considerations. Historically the deck height of a flat car tended to be very closely related to the height of the upper flange of the center sill. This height was itself established by the height of the cap of the draft pocket. The size of the draft pocket was standardised on the basis of the coupler chosen, and the allowable heights for the coupler knuckle. The deck height usually worked out to about 40 or 41 inches above top of rail. For some time auto rack cars were designed to a 19 ft height limit. To maximise the internal loading space, it has been considered desirable to lower the main deck as far as possible, particularly in tri-level cars. Since the lading is relatively light, the trucks have tended to be light as well, such as 70 ton trucks, as opposed to 100, 110 or 125 ton trucks for coal, ore, or grain cars at 263,000, 286,000 or 315,000 lbs. Since the American Association of Railroads (AAR) specifies a minimum clearance of 5″ above the wheels, the combination of low deck height, deck clearance, and minimum wheel height set an effective upper limit on the spring travel, and reserve spring travel range available. If softer springs are used, the remaining room for spring travel below the decks may well not be sufficient to provide the desired reserve height. In consequence, the present inventor proposes, contrary to lowering the main deck, that the main deck be higher than 42 inches to allow for more spring travel.
As noted above, many previous auto rack cars have been built to a 19 ft height. Another major trend in recent years has been the advent of “double stack” intermodal container cars capable of carrying two shipping containers stacked one above the other in a well or to other freight cars falling within the 20 ft 2 in. height limit of AAR plate F. Many main lines have track clearance profiles that can accommodate double stack cars. Consequently, it is now possible to use auto rack cars built to the higher profile of the double stack intermodal container cars. The present inventor has chosen to increase the height of the car generally to provide both a suitable internal height for the lading, and to permit the use of softer springs.
While decreasing the primary vertical bounce natural frequency appears to be advantageous for auto rack rail road cars generally, including single car unit rail road cars, articulated auto rack cars may also benefit not only from adding ballast, but from adding ballast preferentially to the end units near the coupler end trucks. As explained more fully in the description below, the interior trucks of articulated cars tend to be more heavily burdened than the end trucks, primarily because the interior trucks share loads from two adjacent car units, while the couple end trucks only carry loads from one end of one car. There are a number of reasons why it would be advantageous to even out this loading so that the trucks have roughly similar vertical bounce frequencies.
Three piece trucks currently in use tend to use friction dampers, sometimes assisted by hydraulic dampers such as can be mounted, for example, in the spring set. Friction damping has most typically been provided by using spring loaded blocks, or snubbers, mounted with the spring set, with the friction surface bearing against a mating friction surface of the columns of the side frames, or, if the snubber is mounted to the side frame, then the friction surface is mounted on the face of the truck bolster. There are a number of ways to do this. In some instances, as shown at p. 847 of the 1984 Car & Locomotive Cyclopedia lateral springs are housed in the end of the truck bolster, the lateral springs pushing horizontally outward on steel shoes that bear on the vertical faces of the side columns of the side frames. This provides roughly constant friction (subject to the wear of the friction faces), without regard to the degree of compression of the main springs of the suspension.
In another approach, as shown at p. 715 of the 1997 Car & Locomotive Cyclopedia, one of the forward springs in the main spring group, and one of the rearward springs in the main spring group bears upon the underside, or short side of a wedge. One of the long sides, typically an hypotenuse of a wedge, engages a notch, or seat, formed near the outboard end of the truck bolster, and the third side has the friction face that abuts, and bears against, the friction face of the side column (either front or rear, as the case may be), of the side frame. The action of this pair of wedges then provides damping of the various truck motions. In this type of truck the friction force varies directly with the compression of the springs, and increases and decreases as the truck flexes. In the vertical bounce condition, both friction surfaces work in the same direction. In the warping direction (when one wheel rises or falls relative to the other wheel on the same side, thus causing the side frame to pivot about the truck bolster) the friction wedges work in opposite directions against the restoring force of the springs.
The “hunting” phenomenon has been noted above. Hunting generally occurs on tangent (i.e., straight) track as rail car speed increases. It is desirable for the hunting threshold to occur at a speed that is above the operating speed range of the rail car. During hunting the side frames tend to want to rotate about a vertical axis to a non-perpendicular angular orientation relative to the truck bolster sometimes called “parallelogramming”. This will tend to cause lateral deflection of the spring group, and will tend to generate a squeezing force on opposite diagonal sides of the wedges, causing them to tend to bear against the side frame columns. This diagonal action will tend to generate a restoring moment working against the angular deflection. The moment arm of this restoring force is proportional to half the width of the wedge, since half of the friction plate lies to either side of the centreline of the side frame. This tends to be a relatively weak moment connection, and the wedge, even if wider than normal, tends to be positioned over a single spring in the spring group.
Typically, for a truck of fixed wheelbase length, there is a trade-off between wheel load equalisation and resistance to hunting. Where a car is used for carrying high density commodities at low speeds, there may tend to be a higher emphasis on maintaining wheel load equalisation. Where a car is light, and operates at high speed there will be a greater emphasis on avoiding hunting. In general the parallelogram deformation of the truck in hunting is deterred by making the truck laterally more stiff. Another method is to use a transom, typically in the form of a channel running from between the side frames below the spring baskets.
One way to raise the hunting threshold is to employ a truck having a longer wheelbase, or one whose length is proportionately great relative to it width. For example, at present two axle truck wheelbases may range from about 5′-3″ to 6′ -0″. However, the standard North America track gauge is 4′-8½″, giving a wheelbase to track width ratio possibly as small as 1.12. At 6′-0″ the ratio is roughly 1.27. It would be preferable to employ a wheelbase having a longer aspect ratio relative to the track gauge. As described herein, one aspect of the present invention employs a truck with a longer wheelbase, preferably about 86 inches, giving a ratio of 1.52. This increase in wheelbase length may tend also to be benign in terms of wheel loading equalisation.
Another way to raise the hunting threshold is to increase the parallelogram stiffness between the bolster and the side frames. It is possible, as described herein, to employ two wedges, of comparable size to those previously used, the two wedges being placed side by side and each supported by a different spring, or being the outer two wedges in a three deep spring group, to give a larger moment arm to the restoring force and to the damping associated with that force.